Back to EveryPatent.com
United States Patent |
5,310,326
|
Gui
,   et al.
|
May 10, 1994
|
Rotary compressor with improved bore configuration and lubrication system
Abstract
A rotary compressor, such as a sliding vane compressor, comprises a housing
having a bore, a rotor assembly operatively arranged to rotate within the
bore, vanes operatively arranged at the rotor assembly to move linearly
relative to the rotor assembly and to form, together with the housing and
rotor assembly, variable chambers, and a valve assembly arranged in the
housing. The bore has a configuration divided into a expansion region of
elliptical shape, a circular transition region, a polynomial-shaped
compression region and a circular sealing region. The lubricant seal is
composed of an oil injection port and oil grooves for trapping and
transporting the oil. The valve assembly is provided in a recess portion
of the housing and is configured as a thin blade covering each discharge
port which has been relieved to minimize undesirably large difference of
the pressure forces between the two sides of the blade. Oil grooves are
provided in the rotor assembly and in end cap assemblies. In addition, the
area seal and lubricant seal can be utilized in other types of compressors
such as a rolling piston-type rotary compressor.
Inventors:
|
Gui; Fulin (Rockledge, FL);
Scaringe; Robert P. (Rockledge, FL)
|
Assignee:
|
Mainstream Engineering Corporation (Rockledge, FL)
|
Appl. No.:
|
944321 |
Filed:
|
September 14, 1992 |
Current U.S. Class: |
418/76; 418/77; 418/99; 418/150; 418/152; 418/179; 418/189; 418/236; 418/259; 418/270; 418/DIG.1 |
Intern'l Class: |
F04C 018/344; F04C 029/02 |
Field of Search: |
418/76,77,99,150,152,179,189,236,238,259,270,DIG. 1
|
References Cited
U.S. Patent Documents
2057381 | Oct., 1936 | Kenney et al. | 418/150.
|
2491351 | Dec., 1949 | Zeitlin | 418/150.
|
3121528 | Feb., 1964 | Rhodes | 418/152.
|
3809511 | May., 1974 | Linder et al. | 418/259.
|
3820924 | Jun., 1974 | Cassidy | 418/DIG.
|
4299097 | Nov., 1981 | Shank et al. | 418/150.
|
4507065 | Mar., 1985 | Shibuya et al. | 418/99.
|
5049052 | Sep., 1991 | Aihara | 418/179.
|
Foreign Patent Documents |
2725238 | Dec., 1977 | DE | 418/259.
|
55-112893 | Sep., 1980 | JP | 428/189.
|
58-62398 | Apr., 1983 | JP | 418/236.
|
2119692 | May., 1990 | JP | 418/259.
|
3242490 | Oct., 1991 | JP | 418/77.
|
Primary Examiner: Vrablik; John J.
Attorney, Agent or Firm: Evenson, McKeown, Edwards & Lenahan
Claims
We claim:
1. A housing for a rotary compressor having a bore with a configuration
made by a process in which a machine is controlled by and operated in
terms of data obtained with a Fortran program as follows
______________________________________
c initial data setup
c write(*,*) `This progrom is to determine the profile of
c write(*,*) `the bore. please input the r.sub.-- rotor now`
c read(*,*) r.sub.-- rotor
c write(*,*) `please input the length the vane sticks out`
c read (*,*) lv
write(*,*) `please input the endmill radius (inch) now`
read(*,*) rm
c rm =0.0
rbmin=r.sub.-- rotor + .0005
rmin =rbmin-rm
rbmax=r.sub.-- rotor+lv
rmax=rbmax-rm
omg=2.*Pi*6500/60.
a1=rmin
b1=rmax
theta1m=100.0
a2=rmax
b2=rmin
theta2m=120.0
n-2000
write(*,*) `please input the number of points`
read(*,*) n
c set theta dimension
do 100 i=0,n+1
dtheta=360./n
theta(i)=dtheta*(i-1)
c determine the radius as a function of theta
c Region I, the seal region
if (theta(i).ge.b) goto 20
r(i)=rmin
C Region II the expansion region 1
if (theta(i).le.a) goto 100
theta1=(theta(i)-10.)/theta1m *pi/2.
r(i)=radius(a1,b1,theta1)
goto 100
c Region III, the transient region 2
20 r(i)=rmax
if (theta(i).le.c ) goto 100
c Region IV, the compression region
theta2=(theta(i)-c)/theta2m *Pi/2.
r(i)=radius(a2,b2,theta2)
goto 100
100 continue
do 200 i =1,n+1
vt=omg*r(i)*0.0254
v=omg*(r(i)-r(i-1))*0.0254
acel=omg*omg*(r(i)-0.3)*0.0254/9.81
acelratio=(r(i+1)+r(i-1)-2.*r(i))*pi**2/(r(i)-0.3)
c write(*,*)i, theta(i),r(i)
x= r(i)*sin(theta(i)*PI/180.)
y= r(i)*cos(theta(i)*Pi/180.)
write(20,1000)theta(i),r(i), x,y
write(30,2000) x,y
c if (amod(i-1,5).ne.0) goto 200
c write (20,1002)theta(i),r(i),vt,acel,v,acelratio
200 continue
c do 300 i=1,100
c theta(i)=360./99*(i-1)
c x=0.7123*sin(theta(i)*pi/180.)
c y=0.7123*cos(theta(i)*pi/180.)
c300 write(20,2000) x,y
c to calculate the circumference of the profile
cL3=0.0
cL2=0.0
r(n+1)=r(1)
theta(n+1)=theta(1)+360.
do 400 i=2,n+1
cL2=cL2+r(i)*( (theta(i)-theta(i-1)) *pi/180.)
cL3=cL3+r(i-1)*( (theta(i)-theta(I-1)) *pi/180.
400 continue
write(*, *) `circomference=`,cL2,cL3,`de=`,cL2/Pi
1000 format(1x,`.vertline.`,f6.1,`.vertline.`,f8.4,2(`.vertline.`,f8.4),`
.vertline.` )
1002 format(1x,f6.4,4(`,`,f13.4),`,`,f9.5)
2000 format(1x,`X`,f7.4,`Y`,f7.4)
Stop
end
function radius(a,b,theta)
radius=a/sqrt( 1.-(1.-a*a/(b*b))*sin(theta)*sin(theta))
end,
wherein
r.sub.-- rotor -- rotor radius
rmin, rmax - min. and max. radius
a -- the angle at which the first ellipse starts
b -- the angle at which the first ellipse ends the
large circle are begins
c -- the angle at which the large circle ends and the
second ellipse begins
theta -- angle variable (degree)
r -- radius variable (inch)
x,y -- point coordinates
omg -- rotating velocity
Vt -- tangential velocity
V -- Radial velocity
acel -- the tangential acceleration of the vane
acelratio -- the ratio of the radial accel. to the
tangential accl
______________________________________
such that the bore has a configuration divided into an expansion region of
elliptical shape, a circular transition region, a polynomial-shaped
compression region, and a circular sealing region.
2. A rotary compressor comprising a housing having a bore; and vanes
operatively arranged at the rotor assembly to move linearly relative to
the rotor assembly and to form, together with the housing and rotor
assembly, variably chambers, wherein the bore has a composite
configuration divided into an expansion region of elliptical shape, a
circular transition region, a polynomial-shaped compression region and a
circular sealing region, and wherein the bore has a configuration made by
a process in which a machine is computer controlled by and operated in
terms of data obtained with a Fortran program as follows:
______________________________________
c initial data setup
c write(*,*) `This progrom is to determine the profile of
c write(*,*) `the bore. please input the r.sub.-- rotor now`
c read(*,*) r.sub.-- rotor
c write(*,*) `please input the length the vane sticks out`
c read (*,*) lv
write(*,*) `please input the endmill radius (inch) now`
read(*,*) rm
c rm =0.0
rbmin=r.sub.-- rotor + .0005
rmin =rbmin-rm
rbmax=r.sub.-- rotor+lv
rmax=rbmax-rm
omg=2.*Pi*6500/60.
a1=rmin
b1=rmax
theta1m=100.0
a2=rmax
b2=rmin
theta2m=120.0
n-2000
write(*,*) `please input the number of points`
read(*,*) n
c set theta dimension
do 100 i=0,n+1
dtheta=360./n
theta(i)=dtheta*(i-1)
c determine the radius as a function of theta
c Region I, the seal region
if (theta(i).ge.b) goto 20
r(i)=rmin
C Region II the expansion region 1
if (theta(i).le.a) goto 100
theta1=(theta(i)-10.)/theta1m *pi/2.
r(i)=radius(a1,b1,theta1)
goto 100
c Region III, the transient region 2
20 r(i)=rmax
if (theta(i).le.c ) goto 100
c Region IV, the compression region
theta2=(theta(i)-c)/theta2m *Pi/2.
r(i)=radius(a2,b2,theta2)
goto 100
100 continue
do 200 i =1,n+1
vt=omg*r(i)*0.0254
v=omg*(r(i)-r(i-1))*0.0254
acel=omg*omg*(r(i)-0.3)*0.0254/9.81
acelratio=(r(i+1)+r(i-1)-2.*r(i))*pi**2/(r(i)-0.3)
c write(*,*)i, theta(i),r(i)
x= r(i)*sin(theta(i)*PI/180.)
y= r(i)*cos(theta(i)*Pi/180.)
write(20,1000)theta(i),r(i), x,y
write(30,2000) x,y
c if (amod(i-1,5).ne.0) goto 200
c write (20,1002)theta(i),r(i),vt,acel,v,acelratio
200 continue
c do 300 i=1,100
c theta(i)=360./99*(i-1)
c x=0.7123*sin(theta(i)*pi/180.)
c y=0.7123*cos(theta(i)*pi/180.)
c300 write(20,2000) x,y
c to calculate the circumference of the profile
cL3=0.0
cL2=0.0
r(n+1)=r(1)
theta(n+1)=theta(1)+360.
do 400 i=2,n+1
cL2=cL2+r(i)*( (theta(i)-theta(i-1)) *pi/180.)
cL3=cL3+r(i-1)*( (theta(i)-theta(I-1)) *pi/180.
400 continue
write (*, *) `circomference=`,cL2,cL3,`de=`,cL2/Pi
1000 format(1x,`.vertline.`,f6.1,`.vertline.`,f8.4,2(`.vertline.`,f8.4),`
.vertline. `)
1002 format(1x,f6.4,4(`,`,f13.4),`,`,f9.5)
2000 format(1x,`X`,f7.4,`Y`,f7.4)
Stop
end
function radius(a,b,theta)
radius=a/sqrt( 1.-(1.-a*a/(b*b))*sin(theta)*sin(theta))
end,
wherein
r.sub.-- rotor -- rotor radius
rmin, rmax - min. and max. radius
a -- the angle at which the first ellipse starts
b -- the angle at which the first ellipse ends the
large circle are begins
c -- the angle at which the large circle ends and the
second ellipse begins
theta -- angle variable (degree)
r -- radius variable (inch)
x,y -- point coordinates
omg -- rotating velocity
Vt -- tangential velocity
V -- Radial velocity
acel -- the tangential acceleration of the vane
acelratio -- the ratio of the radial accel. to the
tangential accl.
______________________________________
3. The rotary compressor according to claim 2, wherein a curve to each of
the regions is tangential at a point of conjunction with adjoining
regions.
4. The rotary compressor according to claim 2, wherein the expansion region
is defined by a crank angle of 100.degree., at an end of which adjoining
the transition region the associated vane is fully extended.
5. The rotary compressor according to claim 2, wherein the vanes are
arranged in slots in the rotor assembly with a position and angle of the
slots sized to hold vanes of longer length without substantially
decreasing strength of the rotor assembly.
6. The rotary compressor according to claim 2, wherein the valve assembly
is arranged in a recessed area of the housing.
7. The rotary compressor according to claim 2, wherein an axial suction
inlet is arranged at the expansion region.
8. The rotary compressor according to claim 2, wherein the rotor assembly
has at least one oil grove on each end face thereof.
9. The rotary compressor according to claim 2, wherein the sealing region
is configured as an area seal between a higher pressure discharge side and
a lower pressure inlet side.
10. The rotary compressor according to claim 9, wherein the area seal is
defined by the radii of the rotor assembly and the radius of the sealing
region being substantially identical.
11. The rotary compressor according to claim 2, wherein end cap assemblies
are provided at each face of the housing adjacent end faces of the rotor
assembly and include at least one oil injection port and one oil supply
line.
12. The rotary vane machine according to claim 11, wherein at least one oil
groove is provided on the rotor and configured to trap a sufficient amount
of oil, to transport the oil remotely from an injection post and to spread
the oil.
13. The rotary compressor according to claim 2, wherein the vanes are
comprised of self-lubricating material.
14. The rotary compressor according to claim 13, wherein the material is
polyimide.
15. The rotary compressor according to claim 2 further comprising a valve
assembly arranged in a recess of the housing, wherein the valve assembly
comprises at least one thin flexible blade corresponding to at least one
discharge port in the housing and normally covering the at least one
discharge port in the absence of discharge pressure.
16. The rotary compressor according to claim 15, wherein the blade is
spring steel.
17. The rotary compressor according to claim 15, wherein a retainer is
arranged above the at least one flexible blade and is sized and configured
to limit movement of an associated blade away from an associated one of
the at least one discharge port.
18. The rotary compressor according to claim 15, wherein the housing
comprises a crankcase-type oil reservoir.
19. The rotary compressor according to claim 15 wherein an oil reservoir is
connected with the housing, and baffles are arranged in the reservoir so
as to separate liquid from a liquid/gas mixture.
20. The rotary compressor according to claim 15, wherein the at least one
discharge port is radially disposed in close proximity to a circular
sealing region of the bore from the bore surface to a surface of the
recess, and a sealing area at the discharge port for the at least one
blade is relieved to minimize inside and outside pressure differences.
21. The rotary compressor according to claim 20, wherein at least one oil
release groove is arranged along a surface of the bore at an entrance to
the sealing region in proximity to the at least one discharge port so as
to smoothly discharge liquids through the at least one discharge port.
22. The rotary compressor according to claim 20, wherein the at least one
discharge port is substantially tangent to a cylindrical surface of the
rotor assembly.
23. The rotary compressor according to claim 15, wherein oil grooves are
distributed on faces of the rotor assembly and are configured to increase
an oil spreading area.
24. The rotary compressor according to claim 23, wherein each side of the
housing is provided with end cap assemblies adjacent the faces of the
rotor assembly, said end cap assemblies comprising an end disk having an
oil injection port and an end cap having an oil supply groove operatively
associated with the oil injection port.
25. The rotary compressor according to claim 24 wherein a suction inlet is
located in one of the end caps and is configured to fit a curvature of the
bore in an elliptical expansion region of the bore, the suction inlet
beginning proximate a circular sealing region of the bore and angled to
allow a vane gradually to isolate the expansion region from an adjacent
region in the bore.
26. The rotary vane compressor according to claim 24 wherein the end disks
are made of a wear-resistant metal, and the end caps are made of
lightweight material.
Description
BACKGROUND AND SUMMARY OF THE INVENTION
The present invention relates to a rotary compressor and, more
particularly, to a reliable, extremely high pressure ratio, high
efficiency, and lightweight sliding-vane rotary-compressor which has
application in heat pump, air conditioning, and refrigeration
(vapor-compression) applications as well as for the compression of other
fluids such as air, nitrogen, and argon. The compressor can be configured
as a lubrication-free device utilizing self-lubricating materials or as an
oil lubricated device with very high compression ratio capability.
Conventional vane compressors have a relatively low pressure capability,
i.e. somewhere in the range of 65 psi. Moreover, these compressors have
insufficient volumetric and overall thermal efficiency for use in today's
environment where high efficiency components are necessary for heat pumps,
refrigeration equipment and air-conditioning units. Also, the conventional
compressors are not completely compatible with the newer, environmentally
safe refrigerants such as R-134a.
In applications such as aircraft or electronics cooling systems where low
weight, small size, reliability, and efficiency are important criteria, as
well as in refrigerant recovery apparatus where stringent thermal
requirements exist, both the conventional sliding-vane compressors and
other types of compressors such as reciprocating compressors are
unacceptable. The conventional vane compressors do not provide adequate
pressure ratio, and the other types of compressors are too heavy and are
also difficult to obtain a high pressure through one stage. Moreover, the
reciprocating compressors have considerable intake losses (pressure drop)
and large reciprocating acceleration forces that result in rough operation
of the compressor. This kind of compressor also require a relatively large
starting torque and hence a large motor.
It is, therefore, an object of the present invention to develop a
lightweight, high compression ratio rotary compressor to overcome or
eliminate the problems and disadvantages associated with both conventional
vane compressors and the other types of compressors, particularly those
compressors designed for use in compact cooling and refrigerant recovery
application where extreme high pressure ratio exists.
It is another object of the present invention to provide compressors,
particularly sliding-vane rotary compressors, configured adequately to
solve the problems of frictional heat, internal leakage, and low pressure
capacity encountered in conventional rotary compressors.
It is a further object of the present invention to configure a sliding-vane
rotary compressor which maximizes compression ratio, flow rate, and
thermodynamic and volumetric efficiency.
Yet another object of the present invention is to provide a structurally
simple sliding-vane rotary compressor, especially useful in micro-climate
cooling systems and refrigerant recovery systems, which is lightweight,
reliable, adaptable to unusual thermal requirements, smooth running, and
operable with small initial starting torque.
Still a further object of the present invention is to configure a
sliding-vane rotary compressor that is particularly suitable for high
compression ratio of more than 100:1 in one stage.
The foregoing objects and advantages of the present invention have been
achieved with a sliding-vane rotary compressor having an improved bore
geometry, vane slot arrangement, valve assembly, and injected lubrication
sealing effects which permit, by way of example, gas or vapor to be drawn
at low pressure (say, 2.5 psia) and discharged at high pressure (say, 365
psia) in one stage.
The bore geometry of the rotary compressor of the present invention is such
that the curvature used is neither a single circle nor a single ellipse,
as in conventional rotary compressors. In the present invention, the
outline of the bore has been configured, through a computer program, to
have the best mechanical and thermodynamic performance, i.e., to maximize
the inlet flow in the inlet area and maximize the compression ratio in the
compression area, to minimize the dead volume and keep low exhaust
resistance, to minimize the internal leakage and maximize the volume
efficiency, to minimize the frictional heating, and, to increase
thermodynamic efficiency.
A sliding-vane rotary compressor incorporating the features of the present
invention has the following advantages:
1. Much higher pressure capability, up to about 450 psi or more as compared
with conventional vane compressors with pressure capability of only 65
psia;
2. Extremely high compression ratio;
3. Improved volumetric and overall thermal efficiency;
4. Improved internal sealing because of the sealing effect of injected
lubricant resulting in high volumetric flow rates as well as the
aforementioned high pressure capability and high pressure ratios;
5. Low friction resulting in reduced wear, increased life, less frictional
heat, and high efficiency;
6. Compact size; and
7. Compatibility with all refrigerants including the more environmentally
safe refrigerants.
BRIEF DESCRIPTION OF THE DRAWINGS
These and further objects, features and advantages of the present invention
will become more readily apparent from the following detailed description
of a currently preferred embodiment when taken in conjunction with the
accompanying drawings wherein:
FIG. 1 is an elevational, front view of the assembled sliding-vane rotary
compressor, with one side of the end-disk and end cap, and the valve
assembly removed and a portion of the rotor cut away incorporating the
principles of the present invention;
FIG. 2 is a cross-sectional view taken in the direction shown by the arrows
2, 10 in FIG. 1 but with both end caps, the valve assembly and the oil
baffle installed on the compressor;
FIG. 3 is an isolated view of the bore configuration of the compressor
shown in FIG. 1 illustrating the several working regions and bore
geometry;
FIG. 4 is an isolated view of the rotor used in the compressor shown in
FIG. 1;
FIG. 5A is a isolated partial view of the discharge ducts and oil release
groove on a top part of the compressor bore;
FIG. 5B is a cross-sectional view along line 5B--5B of FIG. 5A;
FIG. 6 is a cross-sectional side view of the valve assembly on the
compressor housing;
FIG. 7 is an isolated perspective view of the valve assembly shown in FIG.
6 and exploded to show the two major components thereof;
FIG. 8 is an isolated, perspective exploded view of a pair of an end-disk
and end-cap illustrating the structure of the engraved oil supply line
between the end-disk and end-cap;
FIG. 9 is a schematic view of an embodiment of the compressor of the
present invention having a crankcase as used in a conventional
refrigeration system;
FIG. 10 is similar to FIG. 2, but shows a cross-sectional view of the
compressor with a separate oil reservoir in the direction of arrows 2, 10
in FIG. 1;
FIG. 11 is a cross-sectional front elevational view of a rolling
piston-type rotary compressor employing principles of the present
invention; and
FIG. 12 is a cross-sectional view taken along line 12--12 FIG. 11
illustrating the lubricant injection ports.
DETAILED DESCRIPTION OF THE DRAWINGS
Referring now to the drawings and, in particular, to FIGS. 1 and 2, the
compressor is designated generally by the numeral 10 and comprises a bore
housing 11 having a bore 12 therethrough, a rotor assembly 13 arranged in
the bore 12, and vanes 14. The illustrated embodiment shows a three-vane
arrangement. A combination of the bore 12, the rotor assembly 13, and the
vanes 14 forms three variable chambers X, Y, Z for gas suction,
displacement, and compression. It should be understood, of course, that
different numbers of vanes can be selected depending on the desired
volumetric flow rate and pressure ratio, and also whether an exhaust valve
will be used. The compressor bore housing 11 can be fabricated, for
example, from nodular cast iron, e.g. 100-70-03 (A 536), the rotor
assembly 13 from 4340 alloy, and the vanes 14 from a self-lubricating
polyimide material marketed by DuPont under "VESPEL 211" trademark.
End-caps 15a, 16a and end-disks 15b, 16b are provided on each face of the
compressor housing 11 and are secured thereto in a conventional manner by
way of four respective bolts 17, 18 on each side of the housing. The
end-cap 16a and end-disk 16b are provided, as seen in FIG. 8, with an
axial suction port 19 through which gas to be compressed is drawn into the
compressor 10.
The rotor assembly 13 includes a rotor shaft 30 passing through end-cap 15a
and end-disk 15b. The shaft 30 is supported on a suitable conventional
needle bearing 31 arranged in the end cap 15a. An external seal 32, also
of conventional construction, is provided between the shaft 30 and the
end-cap 15a to prevent leakage of gas and lubricant through the end cap
15a.
As shown in FIG. 2, the top of the compressor 10 is provided with a valve
assembly designated generally by numeral 23 and shown in greater detail in
FIGS. 6 and 7. The valve assembly is secured to the flat top of the
compressor bore housing 11. The valve assembly 23 has two components,
namely a very thin, flexible valve blade 24 in the form of individual
fingers corresponding to the number of discharge ports 26 in the
compressor housing 11 and a retainer 25, which is secured to the housing
11 through conventional screws or bolts, also having the same number of
fingers as the blade 24 to limit upward movement of the valve blade
fingers off the discharge ports 26. The valve blade 24 can be about 0.005
inch thickness to reduce the bending forces for opening and to provide a
rapid dynamic response. Spring steel possesses sufficient elasticity for
this purpose. It will be understood that the rest position of the valve
assembly is shown in solid line in FIG. 6 and also on the right hand side
of FIG. 7 whereas the discharge position is shown by the dashed lines in
FIG. 6.
As shown in FIG. 6, the pressure of the gas inside the housing 11 is
designated P.sub.c and the pressure outside the housing 11 is designated
as P.sub.h. When the compressor 10 is running under rated load, the
outside P.sub.h is approximately constant whereas the inside pressure
P.sub.c varies cyclically. When P.sub.c is larger than P.sub.h, the
stiffness of the fingers of the valve blade 24 will be overcome so as to
push the valve blade 24 upwardly as shown by the dotted lines in FIG. 6 to
permit the compressed gas to flow out of the housing. Since the valve
blade 24 is thin (about 0.005 inch), it is easily lifted to reduce the
discharge pressure (P.sub.h -P.sub.c). The recessed area 22 serves as a
pressure balance for the pressure forces exerted on the top and bottom
sides of the blade 24. This, in turn, reduces the discharge pressure and
results in increased compressor thermal efficiency.
Referring to FIG. 3, the compressor 10 draws in gas on the right hand side
through the side suction port 19 in the end-cap 16a and end-disk 16b to
provide an axial inlet flow into the compressor 10. As the rotor 13
rotates clockwise, as shown by the arrow B in FIG. 1, the gas sucked in
through the port 19 is compressed as it moves to the left hand side of the
compressor 10 through the transition region (II) as shown in FIG. 3 and is
discharged radially through the discharge ports 26 in the housing 11.
All known sliding vane compressors use, however, either a circular bore
with off-set center line or a single ellipse. The performance of the
compressor of the present invention is markedly improved over these known
compressors because of the unique bore configuration which is divided into
four regions; expansion (I), transition (II), compression (III), and seal
(IV) regions which can be summarized as follows:
______________________________________
Region Location Curvature Function
______________________________________
I A-B Ellipse Expansion
II B-D Circle Transition
III D-E Specially Compression
Modified
(Polynomial)
Ellipse
IV E-A Circle Provide large
seal area
______________________________________
Each region uses a different type of bore curvature, from a simple circle
to provide a large sealing area to a modified ellipse to maximize
compression. The bore configuration is made by a CNC machine according to
a copyrighted FORTRAN program owned by applicants' assignee, Mainstream
Engineering Corporation of Rockledge, Fla. This program is machine
implementable to calculate the coordinates of the compressor bore in terms
of selected bore parameters (specifically, rotor radius and vane extending
length), to calculate the velocity and acceleration of the vane, and thus
to provide and control the circumference of the bore profile (or the
center path of a CNC machine end-mill) as represented by the following
source code in which bore profile circumference (or center path) are in
the cartesian coordinate system and in which
______________________________________
r.sub.-- rotor -- rotor radius
rmin, rmax - min. and max. radius
a -- the angle at which the first ellipse starts
b -- the angle at which the first ellipse ends the
large circle are begins
c -- the angle at which the large circle ends and the
second ellipse begins
theta -- angle variable (degree)
r -- radius variable (inch)
x,y -- point coordinates
omg -- rotating velocity
Vt -- tangential velocity
V -- Radial velocity
acel -- the tangential acceleration of the vane
acelratio -- the ratio of the radial accel. to the
tangential accl.
program compbore
common theta(2100) , r(2100)
doubleprecision theta,r
parameter(Pi=3.141593,a=10.,b=110.,c=240.)
open(20,file= `bore.dat` ,status = `unknown`)
open(30,file= `bore2.dat` ,status= `unknown`)
c initial data setup
c write(*,*)`This progrom is to determine the profile of
c write(*,*) ` the bore. please input the r.sub.-- rotor now`
c read(*,*) r.sub.-- rotor
c write(*,*) `please input the length the vane sticks out`
c read(*,*) lv
write(*,*) ` please input the endmill radius (inch) now`
read(*,*)rm
c rm =0.0
rbmin=r.sub.-- rotor + .0005
rmin =rbmin-rm
rbmax=r rotor+lv
rmax=rbmax-rm
omg=2.*Pi*6500/60.
a1=rmin
b1=rmax
thetalm=100.0
a2=rmax
b2=rmin
theta2m=120.0
n-2000
write(*,*) `please input the number of points`
read(*,*) n
c set theta dimension
do 100 i=0,n+1
dtheta=360./n
theta(i)=dtheta*(i-1)
c determine the radius as a function of theta
c Region I, the seal region
if (theta(i).ge.b) goto 20
r(i)=rmin
C Region II the expansion region 1
if (theta(i).le.a) goto 100
thetal=(theta(i)-10.)/thetalm *pi/2.
r(i)=radius (a1,b1,thetal)
goto 100
c Region III, the transient region 2
20 r(i)=rmax
if (theta(i).le.c ) goto 100
c Region IV, the compression region
theta2=(theta(i)-c)/theta2m *Pi/2.
r(i)=radius(a2,b2,theta2)
goto 100
100 continue
do 200 i=1,n+1
vt=omg*r(i)*0.0254
v=omg*(r(i)-r(i-1))*0.0254
acel=omg*omg*(r(i)-0.3)*0.0254/9.81
acelratio=(r(i+1)+r(i-1)-2.*r(i))*pi**2/(r(i)-0.3)
c write(*,*)i,theta(i),r(i)
x= r(i)*sin(theta(i)*PI/180.)
y = r(i)*cos(theta(i)*Pi/180.)
write(20,1000)theta(i),r(i), x,y
write(30,2000) x,y
c if (amod(i-1,5).ne.0) goto 200
c write(20,1002)theta(i),r(i),vt,acel,v,acelratio
200 continue
c do 300 i=1,100
c theta(i)=360./99*(i-1)
c x=0.7123*sin(theta(i)*pi/180.)
c y=0.7123*cos(theta(i)*pi/180.)
c300 write(20,2000) x,y
c to calculate the circumference of the profile
cL3=0.0
cL2=0.0
r(n+1)=r(1)
theta(n+1)=theta(1)+360.
do 400 i=2,n+1
cL2=cL2+r(i)*( (theta(i)-theta(i-1)) *pi/180.)
cL3=cL3+r(i-1)*( (theta(i)-theta(I-1)) *pi/180.
400 continue
write(*, *) `circomference=`, cL2,cL3,` de=`,cL2/Pi
1000 format(1x,`.vertline.`,f6.1,` .vertline.`,f8.4,2(` .vertline.`,f8.4),
` .vertline.`)
1002 format(1x,f6.4,4(`,`,f13.4),`,`,f9.5)
2000 format(1x, `X`,f7.4,`Y`,f7.4)
Stop
end
function radius(a,b,theta)
radius=a/sqrt( 1.-(1.-a*a/(b*b))*sin(theta)*sin(theta)
end
______________________________________
In the inlet or expansion region (I) shown in FIG. 3, the maximization of
the inlet flow rate is accomplished by opening the compression chamber,
i.e. the space between the rotor 13 and the bore configuration 12,
quickly. This region only takes about 100 degrees of crank rotation for
the vane 14 to fully extend. This provides the maximum possible volume in
the transition region (II). The first or inlet region encompasses A-C-C'
since the suction port section is not isolated from the segment until the
trailing-vane 14 of a segment passes point C. The bore curvature 12
changes, however, from a circle to an ellipse at point B. The bore
curvature change, at each region, is also restrained so that the curve
from the previous region and the curve for the next region are tangential
at the point of conjunction. This minimizes accelerations and jerk on the
vanes. The second or transition region (C-C'-D'-D) is about 13% larger
than would be the case if a single circular bore were used, and much
larger than that of an elliptical bore configuration.
The compression region (III) shown in FIG. 3 is an important region because
most of the compression work is accomplished there. Also the vane wear,
heat generation, fluid leakage, and vibration are greatest in the
compression region. All these factors are principally related to the
curvature in this third or compression region. We have also recognized the
importance of carefully considering both mechanical smoothness and
effective thermodynamic performance in determination of the compression
curvature which provides for a smooth movement of the vanes. The radial
acceleration of the vanes varies gently throughout the compression zone.
This reduces vane wear and compressor vibration.
In addition, the compression rate is slow at the beginning (in the area
between the transition region (II) and compression region (III)) and very
fast in the compression region (III). This arrangement results in two
advantages. First, since most thermodynamic compression heat is generated
in the compression region (III), the fast compression and discharge
reduces the time and the area for the hot compressed air to transfer its
heat to the compressor body; hence, the temperature of the compressor is
reduced. Significant heating of the inlet gas would cause the gas to
expand and thereby undesirably reduce the mass flow rate of the compressor
and decrease the maximum compression ratio; the reduction of the
compressor temperature also increases the thermodynamic efficiency.
Second, the fast compression and discharge also reduces the residence time
in the compressor, thereby reducing the internal leakage, and increasing
the volumetric and thermodynamic efficiency. The bore curvature of the
present invention provides, therefore, a significantly improved pressure
capability of this compressor when compared to conventional rotary
compressors.
Another major advantageous feature of the compressor bore is the sealing
zone (IV) E-A which is between the high pressure or discharge side of the
compressor 10 and the low pressure or inlet (suction) side. In contrast
with a line seal in a conventional rotary compressor, the compressor 10
uses an area seal that much more effectively reduces the gas leakage from
high pressure side to low pressure side. The area seal is brought about by
making the radii of both the rotor 13 and the bore 12 in the seal zone as
close to each other as possible within practical manufacturing limits.
As shown in FIG. 5, the discharging holes 26 in the compressor bore housing
11 are arranged radially to avoid exhaust choking and to minimize pressure
drop. Oil release grooves 27, which lead to the discharging holes 26, are
located at the entrance of the contact seal zone. A common compressor
damage problem is liquid knocking which results from compressing a liquid.
Any liquid, such as oil, liquid water, or unvaporized refrigerant, in this
area will be smoothly discharged to the exhaust ports 26 without causing
any knocking due to the oil release grooves 27. Without these grooves 27,
however, liquid which is virtually incompressible in this area would be
forced into the seal contact area causing the rotor torque to increase
tremendously and resulting in the stalling of the compressor or in
undesired displacement of the rotor 13 or vanes 14. Any of these unwanted
effects will damage the compressor or, at the very least, decrease its
life. The oil release grooves 27 allow the compressor 10 to successfully
operate with a liquid volume ratio as high as 50%.
The exhaust ports 26 provide a path for discharging the compressed gas and
use the valve assembly 23 to prevent back flow. They are composed of four
radial holes 26 in the bore housing 11 and utilize the flapper-type valve
assembly 23 on the outside surface of the compressor housing 11. This
arrangement makes a smooth streamlined passage to reduce the flow
resistance during discharge and reduces the dead volume, i.e. the
discharge holes. The discharge holes 26 extend inside the bore as short
and close as possible to the seal region (IV). The cross-sectional flow
area (i.e., the width between the rotor 13 and the bore 12 multiplied by
the length of rotor 13) is just large enough to allow the compressed gas
to exit through this space without choking, thereby reducing the dead
volume to a minimum. The ports 26 have also been sized and configured to
be tangent to the rotor cylinder surface so as to streamline the exhaust
flow and further reduce flow resistance.
The rotor 13 is the only rotating part in the compressor 10. The rotor 13
has slots 28 (FIG. 4) therein to house the sliding vanes 14 and drives the
vanes 14 to displace and compress the gas by making the variable
compression chambers X, Y, Z with the help of the previously discussed
bore curvature 12. It also has oil grooves 29 on its two flat faces for
trapping lubricant and reducing the fluid shear force between the rotor 13
and the bore 12. These oil grooves 29 also supply the lubricant required
for sealing.
The number of the vanes is related to the friction heat generation, seal
capability, and initial compression volume. There is friction between the
vanes 14, the bore 12, and the rotor 13. For a small size compressor, the
unit frictional heat, i.e., the frictional heat generated for a unit of
gases, is high. Excessive friction and frictional heat result in the
reduction of the intake mass flow rate, because of the thermal expansion
of the gas, and the increase of the required unit compression work.
Frictional heating is one of the main concerns in the determination of the
number of vanes. Three vanes have been selected in the currently preferred
embodiment for a refrigeration application because this results in the
largest flow rate and least friction heat compared, say, to five vanes.
Although a fewer number of the vanes does not provide a high leakage
resistance at the tip of the vane, this loss is partially compensated by
the above-mentioned rapid compression. In any event the number of vanes
can be changed without departing from the present invention.
The position and the angle of the slots 28 for the vane 14 shown in FIG. 4
are determined in such a way that longer vanes can be held in the slots
while the rotor still has sufficient strength. This arrangement greatly
increases the vane extending length and hence the volumetric flow rate.
The size, location, and the pattern of the oil grooves 29, which are very
important, are determined in terms of sealing and lubrication. The
end-disks 15b, 16b each have only one oil injection port 37a, 38a (FIG. 8)
located near the seal zone (IV) of the compressor and between the high
pressure side and low pressure side. Oil has to be brought and spread as
much as possible over the whole area of each end-disk which contacts the
rotor face. The oil grooves 29 trap sufficient oil and transport it to the
area far away from the injection port 38a. Two piece-wise grooves between
adjacent vanes are determined to eliminate the possible circumferential
leakage of the gas through the grooves within which is the mixture of gas
and oil. The rising of the grooves 29 (i.e., the increase of the groove
radius) at the tail is to increase the oil spreading area. Experimental
results showed that the seal is very good; a vacuum of thirty inches of
mercury has been reached on the suction side while the discharge side
remains at a pressure above 450 psig.
High machining quality is needed for the rotor 13 because a good sealing
ability relies on a very small clearance between mechanical parts. The
machining tolerance is limited at about 0.00025 inch using presently
available CNC machines.
The vanes 14 are made of a self-lubricating polyimide material which
contains 15% graphite and 15% P.T.F.E. (Teflon). As previously noted, this
material is made by DuPont and marketed under the "VESPEL 211" trademark.
It is lightweight and has a very small frictional coefficient when sliding
on metal. The material was selected based on its compatibility with
refrigerants and lubricants.
Since valve forces are the result of pressure multiplied by surface area, a
large difference between the area exposed to the internal pressure and the
area exposed to the external pressure leads to a large over-pressure
required to open the valve. To reduce this effect, the valve seating area
36 (FIG. 7) is minimized, by relieving the area around the outside of the
seating area 36 with the recessed area 22, so the valve blade 24 contacts
on a small ring area 36, thereby minimizing the required pressure
difference necessary for opening the valve.
The suction inlet 19 is located on one side of the compressor, i.e. the
end-cap 16a in the illustrated embodiment. The shape of the inlet 19 fits
the bore curvature 12 and has maximum area which covers the expansion
cross-sectional area. The inlet 19 starts at point close to the sealing
region (IV) and ends at Point C shown in FIG. 3. A sharp angle of about
45.degree. is provided at the closing point C so that the vane 14 can
isolate the chamber gradually, thereby reducing the noise created by the
periodic suction. The gas is sucked into the compressor 10 axially from
the side through the inlet 19. Because there is no intake valve, a suction
pressure drop does not exist.
On smaller compressors, the rotor 13 is weakened by the vane slots 28.
Thus, deformation of the rotor 13 and loss of parallelism in the rotor
slots 28 can occur. To improve the strength at the root (the radially most
inward portion) of the slot 28 on the rotor 13, the rotor shaft 30 and the
rotor 13 can, for example, be made from one piece of high strength steel,
e.g. 4340 steel alloy. The entire piece is hardened so that a good surface
finish and wear resistance can be gained on both the rotor and shaft
(bearing) surfaces. Larger compressors can, however, utilize a separate
shaft pressed onto the rotor 13 since there is a greater amount of
material and, therefore, greater strength in the root of the vane slots 28
to prevent deformation of the rotor 13.
High thermal efficiency and high pressure ratio are attained via another
main advantageous feature of the invention, namely the lubricant internal
sealing. The lubrication system is composed of oil injection ports 37a,
38a, oil supply grooves 37b, 38b (FIGS. 2 and 8), the oil trapping grooves
29 (FIG. 4), oil supply flow control orifices 42, and an oil-reservoir/oil
separator 39. A working fluid-compatible oil must be used as a lubricant.
For refrigeration systems, refrigerant-compatible oils such as SUNISO 3GS
for R-12, R-22, R-114, R-113, R-500, etc., and Castrol SW68 for R-134a can
be used as the lubricant which plays two roles in this compressor. First,
it lubricates the mechanical moving parts, such as the vanes 14 and
bearings, and, with the help of the oil grooves 29 on the rotor face,
seals the leakage path by viscous effects. This is extremely important
since the leakage path between the discharge ports 26 and the inlet 19 has
a path length of only 0.1 inch and the pressure difference can be above
450 psi. If there were not any way to seal this leakage path, it would
function like a narrow gap nozzle, and a significant quantity of
compressed gas could be injected into the inlet side of the compressor 10,
resulting in a tremendous loss in flow rate and pressure rise capability.
This is particularly critical in small volume compressors.
The leakage rate is directly proportional to the fourth power of the gap
width and inversely proportional to the fluid viscosity for a narrow gap
low Reynolds flow situation. The width is minimized via high quality
manufacturing. The viscosity of the lubricant is about 100 centipoise
while the gas viscosity is only in the order of 0.01 centipoise.
Therefore, the presence of oil in the sealing region results in a
10,000-times increase in viscosity, and thus a 10,000-times decrease in
the leakage rate, if the leakage path is filled with oil. In this way,
excellent sealing is obtained.
No pump is needed for supplying the lubricant to the required lubrication
positions. The lubricant oil is pushed from its reservoir, which could be
either a crankcase as shown in FIG. 9 or a separate
oil-separator/oil-reservoir, into the compressor 10 by the high pressure
produced by the compressor. FIGS. 2 and 8 show that the oil at high
pressure (compressor discharging pressure) is transported to the oil
supply ports 37a, 38a via the two orifices 42, and the two oil-supply
lines 37b, 38b, one on each side of the compressor 10. It can be seen from
FIG. 8 that the oil-supply line 38b is formed by grooves on the end-cap
covered by an end-disk. The orifices 42 inside oil-supply lines 49 (i.e.,
at the bottom of the end-caps) act as a flow-control to assure that the
correct amount of oil is sent to the compressor. The oil is injected into
the compressor 10 from the oil injection ports 37a, 38a, through a pair of
small holes on the end-disks 15b, 16b, and flows in the narrow clearance
between the respective end-disk 15b, 16b and the rotor 13. The oil flows
along this narrow passage under the combined influence of the pressure
gradient and shear forces. Some of the oil is retained in the oil trap
grooves 29 which are located on both faces of the rotor 14 as shown in
FIGS. 1 and 4 and brought to the other part of the end-disk with the
rotation of the rotor 13.
The location of the oil injection ports 37a, 38a is very critical to the
performance of the lubricant seal, and further to the performance of the
entire compressor. Within a small area around a oil injection port, there
is a sharp pressure variation. If the oil injection ports deviate a little
to the high pressure side, much less oil will be injected out; if they are
too close to the low pressure side, part of the area supposed to be sealed
will lack oil while too much oil will be flushed into the compressor. In
both cases, the oil seal will not be sufficiently effective. The position
of the injection port is precisely determined within that critical area so
that the pressure at the location of the oil injection port can be at a
desired value to have the best lubrication and sealing ability.
The oil reservoir has two functions, namely keeping the oil and separating
the oil from the gas and oil mixture. An oil reservoir could be a
crankcase type as seen in FIG. 2 or a separate oil reservoir. The
separation function is accomplished with an oil baffle 40 (FIG. 2). In the
illustrated crankcase reservoir, when the discharged gas and oil mixture
passes through the oil baffle 40, the baffle allows discharged gas to
exit, while the lubricant oil is separated from the gas and trapped in the
crankcase 39 where it accumulates and is reintroduced into the compressor
10. This configuration is more compact and does not involve any fittings.
In an embodiment of the present invention using a detached separator and
reservoir, all the compressor exhaust, i.e. both the discharge gas and
liquid oil, are routed to a separate detached oil-separator/oil-reservoir
which, like the crankcase separator 39, is provided with one or more
baffles to separate the liquid oil from the compressed gas. The oil is
returned to the oil injection ports 37a, 38a (FIG. 10) on the compressor.
It can be cooled prior to returning to the compressor 10. This
configuration allows for an externally cooled reservoir/separator. The oil
inside the separator/reservoir is also under high pressure and is able to
be injected back into the compressor. A separate reservoir has the
advantage of preventing the compressor from contacting the hot compressed
exhaust gas and can therefore run cooler.
The shaft seal 32 is a conventional dual-lip seal and can be made, for
example, of Graphite PTFE (teflon) which has good wear resistance.
Moreover, even if there is some wear, it will not degrade the sealing
ability. A spring and the pressure on the lip will press the seal against
the shaft 30. This assures the reliability of the seal 32 and its long
life. A seal between the compressor and the crankcase is a conventional
O-ring seal. The compressor assembly, i.e. the two end-caps 15a, 16a, the
two end-disks 15b, 16b and the bore housing 11 are fastened together using
the bolts 17, 18 with O-ring seals 45 (FIG. 2) therebetween in a known
manner. This compact and lightweight structure replaces the conventional
method of using large bolts and structure which seals the contact face by
application of large surface forces.
Dual-metal structure, i.e. iron end-disks and aluminum end-caps, is used to
reduce the weight of the compressor for some special needs such as a
lightweight aircraft cooling systems or portable refrigeration systems. A
thin cast-iron disk, which has good wearing resistance, is backed-up by a
thicker, but lightweight, aluminum piece for housing the bearing, seal,
inlet and mounting hardware. Another advantage of the dual metal end-caps
is that the oil supplying grooves, as described above, can be engraved on
the aluminum plate so no external supplying lines needed. The compressor
10 directly picks up the oil through the oil orifice 42 from the bottom of
the crankcase oil reservoir 39.
The compressor bore housing 11 can likewise use two different metals for
construction. Instead of a single cast iron bore housing, a thin cast-iron
sleeve can be used to provide the wear resistance, and an outer aluminum
housing can be used to provide the strength, and mounting surface.
Aside from conventional manufacturing techniques for other parts, the
compressor bore configuration 12 is made by a 5-axis CNC milling machine
which has a resolution of 0.0001 inch, based on the data of the bore
configuration optimized with the FORTRAN program disclosed above. The
end-caps can also be made on the same CNC 5-axis milling machine. All the
precision-matching structures are made in one operation. Only one position
pin 43 (FIG. 1) was used between the bore and the end-cap on each side. A
minimum clearance between the bore and the rotor 13, i.e, the seal zone,
is assured by pressing the bore 11 against the rotor 13 when the
compressor 10 is assembled. Two pins 43, 44 are used to secure each
end-disk 15b, 16b and match the holes 45, 46, respectively, shown in FIG.
8.
By way of illustration only, the overall size of a 1 horsepower compressor
is about 3 inches diameter and 2.7 inches long excluding the crankcase.
The displacement volume with that size compressor is 1.0955 cubic inches
per revolution; 1,095.5 cubic inches per minute at 1000 RPM (0.634
ft.sup.3 /min) and 2,191.0 cubic inches per minute at 2000 RPM (1.268
ft.sup.3 /min). The highest operating pressure is 450 psig, and the lowest
suction pressure is 30 inches Hg.
FIGS. 11 and 12 show another type of generally known compressor, namely a
rolling piston-type compressor, but one that incorporates principles of
the present invention and designated generally by the numeral 60. The
compressor comprises a housing 63 having a bore 64, a crankshaft-like
rotor 61 arranged inside the bore 64, a crankshaft driven rolling piston
62 moving eccentrically inside the bore, a spring-based vane 65 and a face
seal insert 66 between the vane 65 and the rolling piston 62. The rolling
piston 62, the bore 64, and the vane 65 and insert 66 forms two variable
chambers X, Y for gas suction and compression. The rolling piston 62
contacts the bore 64 at a contacting point C which rotates clockwise
around the bore 64. When the contact point C passes the inlet port 68, a
"new" chamber X is produced, and the "old" chamber X turns into chamber Y
for a gradual compression operation.
The face seal insert 66 is provided between the vane 65 and the rolling
piston 62 to form an area seal between the high pressure region and the
low pressure region. The intake or inlet port 68 and the discharge port 69
are located on each side of the vane 65. A valve arrangement 70 is
provided over the discharge port 69 and is constructed to operate
substantially in the same way as valve assembly 23 shown in FIGS. 6 and 7.
The lubrication and face seal system discussed with respect to FIGS. 1, 2,
4 and 8 is also essentially utilized in the rolling piston-type rotary
compressor. The lubricant is injected into the gap between the rolling
piston 62 and the end-caps 15, 16 through oil injection ports 37a, 38a
(FIG. 12) and is trapped in oil grooves 29 (FIG. 11) and further
transported to the other part of the face of the rolling piston.
Although the invention has been described and illustrated in detail, it is
to be clearly understood that the same is by way of illustration and
example, and is not to be taken by way of limitation. The spirit and scope
of the present invention are to be limited only by the terms of the
appended claims.
Top